Refrigerating machine having intermediate-pressure receiver

ABSTRACT

A refrigerating machine including a two-stage compressor, a high-pressure gas cooler, a first expansion valve, an intermediate-pressure receiver, a second expansion valve, and an evaporator is further equipped with an intermediate-pressure refrigerant bypass circuit for bypassing gas refrigerant in the intermediate-pressure receiver to an intermediate-pressure portion of the two-stage compressor, a back flow preventing device provided to the intermediate-pressure refrigerant bypass circuit for preventing back flow of refrigerant from the two-stage compressor to the intermediate-pressure receiver, and a refrigerant-pressure control unit for controlling the pressure of the refrigerant in the intermediate-pressure receiver on the basis of the difference between a specific enthalpy of refrigerant discharged from the high-pressure gas cooler and a predetermined reference enthalpy.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a refrigerating machine having an intermediate-pressure receiver, gas refrigerant in the intermediate pressure receiver being supplied to an intermediate pressure portion of a two-stage compressor.

2. Description of the Related Art

There is known a refrigerating machine in which a two-stage compressor, a high pressure gas cooler for cooling high-pressure gas refrigerant, a first throttling device, an intermediate pressure receiver for adjusting a refrigerant circulating amount, a second throttling device and an evaporator that are generally successively connected to one another to thereby form a closed circuit, an intermediate pressure refrigerant bypass circuit for bypassing intermediate-pressure refrigerant vapor in the intermediate pressure receiver to the intermediate pressure portion of the compressor is equipped, and the high pressure portion is driven under a supercritical state under normal operation (see JP-A-2003-106693). In this type of refrigerating machine, gas refrigerant separated in the intermediate pressure receiver is introduced into the intermediate pressure portion of the two-stage compressor while kept under gas state, and thus the refrigerating machine is set to a so-called two-stage expansion economizer cycle, so that the refrigerant flow in the evaporator is reduced and the compression driving force of the first-stage compressor is reduced and the pressure loss in the evaporator is reduced. Therefore, the performance of the refrigeration cycle can be enhanced.

However, in the conventional two-stage expansion economizer cycle, for example when only liquid-phase refrigerant exists in the intermediate pressure receiver due to an external temperature, a load condition or the like, a part of the liquid-phase refrigerant to be introduced into the evaporator is introduced into the intermediate pressure portion of the two-stage compressor, so that the compression efficiency is reduced and also the compressor is damaged or the like by liquid back.

SUMMARY OF THE INVENTION

The present invention has been implemented in view of the foregoing situation, and an object of the present invention is to provide a refrigerating machine that can keep the optimum performance in accordance with an external temperature, a load condition or the like.

In order to attain the above object, according to an aspect of the present invention, there is provided a refrigerating machine comprising a two-stage compressor, a high-pressure gas cooler for cooling high-pressure gas refrigerant discharged from the two-stage compressor, a first throttling device for expanding the gas refrigerant from the high-pressure gas cooler, an intermediate-pressure receiver for adjusting a refrigerant circulating amount, a second throttling device for expanding the refrigerant from the intermediate-pressure receiver and an evaporator that are successively connected to one another to form a closed refrigerant circuit, which is further equipped with an intermediate-pressure refrigerant bypass circuit for bypassing gas refrigerant in the intermediate-pressure receiver to an intermediate-pressure portion of the two-stage compressor, a back flow preventing device that is provided to the intermediate-pressure refrigerant bypass circuit and prevents back flow of refrigerant from the two-stage compressor to the intermediate-pressure receiver, and a refrigerant-pressure control unit for controlling the pressure of the refrigerant in the intermediate-pressure receiver on the basis of the difference between a specific enthalpy of refrigerant discharged from the high-pressure gas cooler and a predetermined reference enthalpy.

In the above-described refrigerating machine, under normal operation, a high pressure portion of the refrigerating machine operates under a supercritical state.

According to another aspect of the present invention, there is provided a refrigerating machine comprises: an outdoor unit including a compressor having an intermediate pressure portion into which intermediate-pressure refrigerant having intermediate pressure between refrigerant pressure at a suction port of the compressor and refrigerant pressure at a discharge port of the compressor can be introduced, and an outdoor heat exchanger serving as a heat-source side heat exchanger; a plurality of indoor units each including an indoor heat exchanger serving as a using side heat exchanger, the plural indoor units carrying out one of cooling operation or heating operation at the same time or carrying out a mixing operation including cooling operation and heating operation at the same time; an inter-unit pipe for connecting the outdoor unit and each of the indoor units to each other, the inter-unit pipe comprising a high-pressure pipe connected to the refrigerant discharge pipe, a low-pressure pipe connected to the refrigerant suction pipe and an intermediate pressure pipe connected to the other end of the outdoor heat exchanger; an intermediate-pressure receiver that is disposed between the heat-source side heat exchanger and the using side heat exchanger and separates gas/liquid mixture refrigerant discharged from any one of the heat-source heat exchanger and the using side heat-exchanger into gas refrigerant and liquid refrigerant and supplying the gas refrigerant to the intermediate-pressure portion of the compressor; a back flow preventing device that is provided between the intermediate-pressure receiver and the intermediate-pressure portion of the compressor and prevents back flow of the gas refrigerant from the compressor to the intermediate-pressure receiver; and a refrigerant-pressure control unit for controlling the pressure of the refrigerant in the intermediate-pressure receiver on the basis of the difference between a predetermined reference enthalpy and a specific enthalpy of refrigerant discharged from any one of the heat-side heat exchanger and the using side heat exchanger.

In the above-described refrigerating machine, under normal operation, a high pressure portion of the refrigerating machine operates under a supercritical state.

According to the refrigerating machine of the present invention, when the specific enthalpy of the refrigerant at the exit of the high-pressure gas cooler (i.e., the refrigerant discharged from the high-pressure gas cooler) is increased due to increase of the external temperature, variation of the load or the like, the two-stage expansion economizer cycle is formed. On the other hand, when the specific enthalpy of the refrigerant at the exit of the high-pressure gas cooler is reduced due to reduction of the external temperature, variation of the load or the like, one-stage expansion cycle is formed. Therefore, the optimum performance can be kept with a simple construction.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a refrigerant circuit diagram showing a refrigerating machine according to an embodiment of the present invention;

FIG. 2 is a pressure-enthalpy (ph) diagram of a refrigeration cycle;

FIG. 3 is a diagram showing a control flow; and

FIG. 4 is a refrigerant circuit diagram of another embodiment.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

A preferred embodiment according to the present invention will be described hereunder with reference to the accompanying drawings.

FIG. 1 is a refrigerant circuit diagram showing an embodiment of the present invention. A refrigerating machine 30 is equipped with a two-stage compressor 1, a high-pressure gas cooler 3 for cooling high-pressure gas refrigerant, a first throttling device 5, an intermediate pressure receiver 7 for adjusting a refrigerant circulating amount, a second throttling device 9 and an evaporator 11, and these parts are successively connected to one another to thereby form a closed circuit. The first throttling device 5 and the second throttling device 9 are designed so that the opening degree of the restriction thereof is variable. By varying the degree of the restriction, the pressure can be reduced and a larger amount of gas refrigerant is generated until it reaches the intermediate pressure receiver 7, and under this state the refrigerant is supplied into the intermediate pressure receiver 7, whereby the separation coefficient in the intermediate pressure receiver 7 can be varied.

The two-stage compressor 1 comprises a first-stage compressing portion 1A and a second-stage compressing portion 1B. The intermediate point (an intermediate pressure portion 1C) between the first-stage compressing portion 1A and the second-stage compressing portion 1B is connected to the upper portion of an intermediate pressure receiver 7 by an intermediate pressure refrigerant bypass circuit 13 for bypassing the intermediate pressure refrigerant vapor in the intermediate pressure receiver 7 to the intermediate pressure portion 1C of the compressor 1, and the intermediate pressure refrigerant bypass circuit 13 is provided with a check valve (back flow preventing device) 15 having the function of preventing back flow of refrigerant vapor from the compressor 1 to the intermediate pressure receiver 7. the back flow preventing device is not limited to the check valve 15, and an opening/closing valve or the like may be used, for example.

Carbon dioxide refrigerant with which the high-pressure side is set to a supercritical state under normal operation is sealingly filled in the above-described refrigerant circuit. Ethylene, diborane, ethane, nitrogen oxide or the like may be used as refrigerant with which the high-pressure side is operated under supercritical pressure.

In this construction, a refrigerant temperature sensor 40 is secured to the exit of the high-pressure gas cooler 3, an evaporation temperature sensor 41 is secured to the evaporator 11, a suction temperature sensor 42 is secured to the suction side of the two-stage compressor 1, a discharge temperature sensor 43 is secured to the discharge side of the two-stage compressor 1, and an intermediate pressure temperature sensor 44 is secured to the intermediate pressure receiver 7. Furthermore, the respective sensors 40 to 44, the first throttling device 5 and the second throttling device 9 are connected to a controller 45.

In this construction, the controller 45 executes the following control.

That is, when the specific enthalpy of refrigerant at the exit of the high-pressure gas cooler 3 is not larger than the enthalpy of saturated liquid corresponding to the pressure of the intermediate pressure portion of the compressor under one-stage expansion, at least one of the first throttling device 5 and the second throttling device 9 is controlled so that the pressure of the intermediate pressure receiver 7 is lower than the pressure of the intermediate pressure portion of the compressor. For example, at least one of the first and second throttling devices is controlled so that the valve opening degree of the first throttling device 5 is “small” and the valve opening degree of the second throttling device 9 is “large”. Furthermore, when the specific enthalpy of the refrigerant at the exit of the high-pressure gas cooler 3 is larger than the enthalpy of saturated liquid corresponding to the pressure of the intermediate pressure portion of the compressor, at least one of the first throttling device 5 and the second throttling device 9 is controlled so that the pressure of the intermediate pressure receiver 7 is lower than the pressure of saturated liquid whose enthalpy is substantially equal to the specific enthalpy of the refrigerant at the exit of the high-pressure gas cooler 3 and also higher than the pressure of the intermediate pressure portion of the compressor. For example, at least one of the throttling devices 5 and 9 is controlled so that the valve opening degree of the first throttling device 5 is “large” and the valve opening degree of the second throttling device 9 is “small”.

FIG. 2 is a pressure-enthalpy (ph) diagram of a refrigeration cycle containing the two-stage compressor, and the high-pressure side is operated under the supercritical state.

In FIG. 2, the pressure “P1” corresponds to the pressure of the intermediate pressure portion of the compressor under one-stage expansion is carried out, and the enthalpy “h1” corresponds to the enthalpy of saturated liquid corresponding to the pressure “P1”. Here, when the external temperature increases or the like, the specific enthalpy “h2” at the exit “E” of the high-pressure gas cooler 3 is larger than the enthalpy “h1” of saturated liquid corresponding to the pressure “P1” of the intermediate pressure portion of the compressor. In this case, at least one of the first throttling device 5 and the second throttling device 9 is controlled so that the pressure “P2” in the intermediate pressure receiver 7 (“F” in FIG. 2) is lower than the pressure “P3” of saturated liquid whose enthalpy is substantially equal to the specific enthalpy “h2” of the refrigerant at exit of the high-pressure gas cooler 3, and also higher than the pressure “P1” of the intermediate pressure portion of the compressor.

Specifically, at least one of the throttling devices 5 and 7 is controlled so that the valve opening degree of the first throttling device 5 is “large” and the valve opening degree of the second throttling device 9 is “small”, for example.

Here, “A” represents the suction state of the first-stage compressing portion 1A, “B” represents the discharge state of the first-stage compressing portion 1A, “C” represents the suction state of the second-stage compressing portion 1B and “D” represents the discharge state of the second-stage compressing portion 1B. The refrigerant discharged from the compressor 1 is passed through the high-pressure gas cooler 3 and circulated to be cooled. “E” represents the exit of the high-pressure gas cooler 3, that is, the entrance of the first restricting deice 5, and “F” represents the exit of the first throttling device 5. Under this state, the refrigerant is two-phase mixture of gas/liquid. The ratio between gas and liquid corresponds to the ratio between the length of a line segment (gas) from “F” to “G” and the length of a line segment (liquid) from “F” to “I”.

This refrigerant enters the intermediate pressure receiver 7 under the two-phase mixture state. The gas refrigerant separated in the intermediate pressure receiver 7 is controlled so that the pressure “P2” of the intermediate pressure receiver 7 is higher than the pressure “P1” of the intermediate pressure portion of the compressor, and thus the refrigerant is passed through the check valve 15 and then introduced into the intermediate pressure portion 1C of the compressor 1, that is, the intermediate point between the first-stage compressing portion 1A and the second-stage compressing portion 1B. “I” represents the exit state of the intermediate pressure receiver 7, and the refrigerant passed through this exit reaches the suction port of the second-stage compressing portion 1B of “C” to be compressed in the second-stage compressing portion 1B.

The liquid refrigerant separated in the intermediate pressure receiver 7 reaches the second throttling device 9. “G” represents the exit of the intermediate pressure receiver 7, that is, the entrance of the second throttling device 9, “H” represents the exit of the second throttling device 9, and “A” represents the exit of the evaporator 11, and also the suction port of the first-stage compressing portion 1A. The liquid refrigerant entering the evaporator 11 is evaporated while absorbing heat, and the gas-phase refrigerant is returned to the suction port of the first-stage compressing portion 1A.

In the above construction, even when the gas refrigerant separated in the intermediate pressure receiver 7 is circulated to the evaporator 11, it cannot be used for cooling. Accordingly, when the gas refrigerant concerned is returned to the suction port of the first-stage compressing portion 1A, the compression efficiency is lowered.

This construction is a so-called two-stage expansion economizer cycle. The gas refrigerant separated in the intermediate pressure receiver 7 is introduced into the intermediate pressure portion 1C of the two-stage compressor 1, so that the refrigerant flow amount in the evaporator 11 is reduced and the compression driving force of the first-stage compressing portion 1A is reduced. Therefore, the pressure loss in the evaporator 11 is reduced, and thus the performance of the refrigeration cycle can be enhanced. In this construction, particularly, carbon dioxide refrigerant is sealingly filled in the refrigerant circuit. Therefore, in the ratio between the gas refrigerant and the liquid refrigerant separated in the intermediate pressure receiver 7, an amount of the gas component (the line segment between “F” and “G”) is higher than Freon (chlorofluorocarbon) refrigerant, and a larger amount of the gas component is introduced into the intermediate pressure portion 1C of the compressor 1, thereby achieving higher performance.

When the external temperature is reduced or the like, the exit state of the high-pressure gas cooler 3 moves to “E1”. The specific enthalpy “h3” of “E1” is smaller than the enthalpy “h1” of the saturated liquid corresponding to the pressure “P1” of the intermediate pressure portion of the compressor, and under this state, only liquid-phase refrigerant exists in the intermediate pressure receiver 7 (“F1”) while no gas refrigerant exists in the intermediate pressure receiver 7.

In this case, at least one of the first throttling device 5 and the second throttling device 9 is controlled so that the pressure “P4” of the intermediate pressure receiver 7 is lower than the pressure “P1” of the intermediate pressure portion of the compressor. For example, at least one of the throttling devices 5 and 9 is controlled so that the valve opening degree of the first throttling device 5 is “small” and the valve opening degree of the second throttling device 9 is “large”. When the pressure “P4” of the intermediate pressure receiver 7 is lower than the pressure “P1”, the check valve 15 of FIG. 1 functions, and the intercommunication between the intermediate pressure receiver 7 and the intermediate pressure portion 1C of the compressor 1 is cut off. All the liquid-phase refrigerant in the intermediate pressure receiver 7 is passed through the evaporator 11 and then introduced into the first-stage compressing portion 1A of the two-stage compressor 1.

When the above operation is considered in FIG. 2, “A” represents the suction of the first-stage compressing portion 1A, “B1” represents the discharge of the first-stage compressing portion 1A and “D1” represents the discharge of the second-stage compressing portion 1B. The refrigerant discharged from the compressor 1 is passed through the high-pressure gas cooler 3 and circulated to be cooled. As described above, “E1” represents the exit of the high-pressure gas cooler 3, that is, the entrance of the first throttling device 5, and “F1” represents the exit of the first throttling device 5. Under this state, only liquid-phase refrigerant exists at “F1”.

All the liquid refrigerant concerned reaches the second throttling device 9. “H1” represents the exit of the second throttling device 9, and “A” represents the exit of the evaporator 11 and also the suction port of the first-stage compressing portion 1A as described above. The liquid refrigerant entering the evaporator 11 is evaporated while absorbing heat, and the gas-phase refrigerant is returned to the suction port of the first-stage compressing portion 1A.

In this embodiment, when the specific enthalpy “h2” of the refrigerant at the exit of the high-pressure gas cooler 3 is larger than the enthalpy “h1” of saturated liquid corresponding to the pressure “P1” of the intermediate pressure portion of the compressor due to increase of the external temperature, variation of load or the like, the two-stage expansion economizer cycle is formed. Conversely, when the specific enthalpy “h3” of the refrigerant at the exit of the high-pressure gas cooler 3 is reduced to be equal to or smaller than the enthalpy “h1” of saturated liquid corresponding to the pressure “P1” of the intermediate pressure portion of the compressor due to reduction of the external temperature, the load variation or the like, one-stage expansion cycle is formed. Therefore, the optimum performance matched with the external temperature, the load variation or the like can be kept with a simple construction.

FIG. 3 shows the control flow of the refrigerating machine. During operation, the evaporation temperature Teva is detected by the evaporation temperature senor 41 (S1) and the suction temperature Tsuc is detected by the suction temperature sensor 42 (S2). Furthermore, the discharge temperature Tdis is detected by the discharge temperature sensor 43 (S3), and the refrigerant temperature Tm in the intermediate pressure receiver 7 is detected by the intermediate pressure temperature sensor 44 (S4). The refrigerant temperature Tout at the exit of the high-pressure gas cooler 3 is detected by the refrigerant temperature sensor 40 (S5). Then, the suction pressure Psuc is calculated from the evaporation temperature Teva (S6), and the high side pressure Ph is calculated from the discharge temperature Tdis (S7), the actual intermediate pressure Pm in the intermediate pressure receiver 7 is calculated from the refrigerant temperature Tm in the intermediate pressure receiver 7 (S8), and the intermediate pressure (=the pressure of the intermediate pressure portion of the compressor when one-stage expansion is carried out) Pm1 as a reference in this control is calculated from the suction pressure Psuc, the suction temperature Tsuc and the high side pressure Ph (S9).

The enthalpy hLiq (“h1”) of saturated liquid corresponding to the intermediate pressure Pm1 is calculated from the intermediate pressure Pm1 (S10), and the specific enthalpy hout (“h2”) at the exit concerned is calculated from the refrigerant temperature Tout at the exit of the high-pressure gas cooler 3 and the high side pressure Ph (S1).

Subsequently, it is judged whether the specific enthalpy hout is larger than the enthalpy hLiq (S12). When the specific enthalpy hout is larger than the enthalpy hLiq (S12:Yes), at least one of the first throttling device 5 and the second throttling device 9 is controlled so as to satisfy the following condition: the intermediate pressure Pm (i.e., the actual pressure in the intermediate-pressure receiver)>the intermediate pressure Pm1 (i.e., the pressure of the intermediate-pressure portion of the compressor) (S13). Specifically, the throttling devices 5 and 9 are controlled so that the valve opening degree of the first throttling device 5 is “large” and the valve opening degree of the second throttling device 9 is “small”, whereby the two-stage expansion economizer cycle is formed.

On the other hand, when the specific enthalpy hout is not larger than the enthalpy hLiq (S12: No), at least one of the first throttling device 5 and the second throttling device 9 is controlled so as to satisfy the following condition: the intermediate pressure Pm (i.e., the actual pressure in the intermediate-pressure receiver)<the intermediate pressure Pm1 (i.e., the pressure of the intermediate-pressure portion of the compressor) (S14). Specifically, the throttling devices are controlled so that the valve opening degree of the first throttling device 5 is “small” and the valve opening degree of the second throttling device 9 is “large”, whereby the one-stage expansion cycle is formed.

The suction pressure Psuc and the high side pressure Ph may be determined by pressure sensors. With respect to the intermediate pressure Pm1, a preset value may be stored in a memory.

FIG. 4 shows another embodiment of the refrigerating machine of the present invention.

A refrigerating machine (air conditioner) 130 can perform both the cooling and heating operation at the same time.

The refrigerating machine 130 comprises an outdoor unit 101 including a two-stage compressor 102, outdoor heat exchangers 103 a and 103 b and outdoor expansion valves 127 a and 127 b, an indoor unit 105 a including an indoor heat exchanger 106a and an indoor expansion valve 118 a, an indoor unit 105 b including an indoor heat exchanger 106 b and an indoor expansion valve 118 b, and a hot water supply (stocking) unit 150 including a hot water stocking heat exchanger 141, a hot water stocking tank 143, a circulating pump 145 and an expansion valve 147.

The outdoor unit 101, the indoor units 105 a and 105 b and the hot water supply unit 150 are connected to one another through inter-unit pipes 110. The refrigerating machine 130 can carry out the cooling operation or the heating operation on the indoor units 105 a and 105 b at the same time or the mixing operation of the cooling operation and the heating operation on the indoor units 105 a and 105 b while operating the hot water supply unit 150.

In the outdoor unit 101, one end of the outdoor heat exchanger 103 a is exclusively connected to the discharge pipe 107 or suction pipe 108 of the compressor 102 through a change-over valve 109 a or change-over valve 109 b. Likewise, one end of the outdoor heat exchanger 103 b is exclusively connected to the discharge pipe 107 or suction pipe 108 of the compressor 102 through a change-over valve 119 a or 119 b. Furthermore, an accumulator 104 is disposed in the suction pipe 108.

The outdoor unit 101 is equipped with an outdoor control device (not shown), and the outdoor control device controls the compressor 102, the outdoor expansion valves 127 a and 127 b, the change-over valves 109 a, 119 a, 109 b and 119 b and the overall refrigerating machine 130. Furthermore, the refrigerating machine 130 is equipped with a temperature sensor S1 for detecting the refrigerant temperature at the entrance of the accumulator 104, temperature sensors S2 for detecting the refrigerant temperature of the indoor heat exchangers 106 a and 106 b, temperature sensors S3 for detecting the refrigerant temperature of the outdoor heat exchangers 103 a and 103 b, and a temperature sensor S4 for detecting the refrigerant temperature at the exit of the compressor 102.

The compressor 102 is a two-stage compressor, and it has a first-stage compressing portion 102A for compressing refrigerant at the low-pressure suction side, and a second-stage compressing portion 102B for compressing refrigerant at the high-pressure discharge side. The compressor 102 is further equipped with an intermediate pressure portion 102M that can introduce refrigerant from the external into the intermediate portion between the first-stage compressing portion 102A and the second-stage compressing portion 102B.

The inter-unit pipe 110 is equipped with a high-pressure pipe (high-pressure gas pipe) 111, a low-pressure pipe (low-pressure gas pipe) 112 and an intermediate-pressure pipe (liquid pipe) 113. The high-pressure pipe 111 is connected to the discharge pipe 107, and the low-pressure pipe 112 is connected to the suction pipe 108. The intermediate pressure pipe 113 is connected to the other ends of the outdoor heat exchangers 103 a and 103 b through the outdoor expansion valves 127 a and 127 b.

An intermediate-pressure receiver (gas/liquid separator) 128 is connected between the intermediate pressure pipe 113 and each of the outdoor expansion valves 127 a and 127 b. The intermediate pressure receiver 128 is roughly equipped with a receiver main body 128A, a vapor exit pipe (serving as an intermediate-pressure refrigerant bypass circuit) 128B, a first entrance/exit pipe 128C and a second entrance/exit pipe 128D, and the vapor exit pipe 128B of the intermediate pressure receiver 128 is connected to the intermediate pressure portion 102M of the compressor 102 so that gas-phase refrigerant is introduced from the vapor exit pipe 128B into the compressor 102. The intermediate pressure receiver 128 is designed as a bidirectional type gas/liquid separator into which refrigerant can be introduced from any side of the outdoor heat exchangers 103 a and 103 b and the indoor heat exchangers 106 a and 106 b.

One ends of the indoor heat exchangers 106 a and 106 b of the indoor units 105 a and 105 b are connected to the high-pressure pipe 111 through discharge side valves 116 a and 116 b, and also connected to the low-pressure pipe 112 through suction side valves 117 a and 117 b. Furthermore, the other ends of the indoor heat exchangers 106 a and 106 b of the indoor units 105 a and 105 b are connected to the intermediate-pressure pipe 113 through the indoor expansion valves 118 a and 118 b. The discharge side valve 116 a and the suction side valve 117 a are controlled so that when one of these valves is opened, the other valve is closed. Likewise, the discharge side valve 116 b and the suction side valve 117 b are also controlled so that when one of these valves is opened, the other valve is closed. Accordingly, one end of each of the indoor heat exchangers 106 a and 106 b is selectively connected to one of the high-pressure pipe 111 and the low-pressure pipe 112 of the inter-unit pipe 110.

The indoor unit 105 a (105 b) is further equipped with an indoor fan 123 a (123 b), a remote controller and an indoor control device. Each indoor fan 123 a (123 b) is disposed in proximity to each indoor heat exchanger 106 a (106 b), and blows air to the indoor heat exchanger 106 a (106 b). Each remote controller is connected to the indoor unit 105 a (105 b), and each indoor unit 105 a (105 b) outputs a cooling or heating operation instruction, a stop instruction, etc. to each indoor control device.

In the hot water supply unit 150, one end of the hot water stocking heat exchanger 141 is connected to the high-pressure pipe 111 through a change-over valve 148, and the other end of the hot water stocking heat exchanger 141 is connected to the intermediate pressure pipe 113 through an expansion valve 147. A water pipe 146 is connected to the hot water stocking heat exchanger 141, and a hot water stocking tank 143 is connected to the water pipe 146 through a circulating pump 145.

In this embodiment, carbon dioxide refrigerant is sealingly filled in the outdoor unit 101, the indoor units 105 a and 105 b and the pipes of the hot water supply unit 150 and the inter-unit pipe 110.

Furthermore, a check valve (back flow preventing device) 151 having the function of preventing back flow of refrigerant vapor from the compressor 102 to the intermediate pressure receiver 128 is provided to the vapor exit pipe 128B of the intermediate pressure receiver 128. The back flow preventing device is not limited to the check valve 151, and for example, an opening/closing valve or the like may be used.

When the refrigerating machine 130 carries out cooling operation or heating operation on the indoor units 105 a and 105 b at the same time or carries out both the cooling operation and the heating operation (i.e., the mixing operation) on the indoor units 105 a and 105 b while operating the hot water supply unit 150, some of the heat exchangers 103, 106, 141 function as a heat-radiation side heat exchanger(s). The gas-phase or liquid-phase component in the refrigerant before the refrigerant enters the intermediate-pressure receiver 128 is varied in accordance with the exit temperature of the heat-radiation side heat exchanger (corresponding to the high-pressure gas cooler 3 of FIG. 1) as described above. When the exit temperature of the heat-radiation side heat exchanger increases or the like, the amount of the gas-phase component of the refrigerant before the refrigerant enters the intermediate-pressure receiver 128 increases, and thus the amount of the gas-phase refrigerant introduced into the intermediate pressure portion 102M of the compressor 102 is increased, so that the efficiency of the refrigeration cycle can be enhanced by the degree corresponding the amount of the gas-phase component which does not contribute to cooling and is not circulated to the low-pressure circuit subsequent to the intermediate-pressure pipe 113.

Particularly, according to the refrigerating machine having the above construction, since carbon dioxide refrigerant is sealingly filled in the refrigerant circuit, in the ratio of the gas-phase component and the liquid-phase component separated in the intermediate-pressure receiver 128, the amount of the gas-phase component separated in the intermediate-pressure receiver 128 is larger as compared with Freon (chlorofluorocarbon) refrigerant, and a large amount of gas-phase component can be introduced into the intermediate pressure portion 102M of the compressor 102, thereby achieving a higher efficiency.

On the other hand, for example when the exit temperature of the heat-radiation side heat exchanger is reduced and thus most of refrigerant in the intermediate-pressure receiver 128 is liquid-phase component, the efficiency of the refrigeration cycle would be rather reduced if the liquid-phase component concerned is introduced to the intermediate-pressure portion 102M of the compressor 102.

In this case, referring to FIG. 2, at least one of the outdoor expansion values 127 a, 127 b, the expansion valve 147 and the indoor expansion valves 118 a, 118 b is controlled so that the pressure “P4” in the intermediate-pressure receiver 128 is lower than the pressure “P1” of the intermediate-pressure portion of the compressor. At this time, the check valve 151 of FIG. 4 functions, and the intercommunication between the intermediate-pressure receiver 128 and the intermediate pressure portion 102M of the compressor 102 is cut off. Therefore, all the liquid-phase refrigerant in the intermediate-pressure receiver 128 is passed through the evaporator, and introduced into the first-stage compressing portion 102A of the two-stage compressor 102.

In other words, when the specific enthalpy “h2” of the refrigerant at the exit of the heat-radiation side heat exchanger increases to be larger than the enthalpy “h1” of saturated liquid corresponding to the pressure “P1” of the intermediate-pressure portion of the compressor due to increase of the external temperature, variation of the load or the like, the two-stage expansion economizer cycle is formed. Conversely, when the specific enthalpy “h3” of the refrigerant at the exit of the heat-radiation side heat exchanger is reduced to be equal to or smaller than the enthalpy “h1” of saturated liquid corresponding to the pressure “P1” of the intermediate-pressure portion of the compressor due to reduction in external temperature or the like, one-stage expansion cycle is formed. Accordingly, the optimum performance which is matched with the external temperature, the load condition or the like can be kept with a simple construction. Here, the “external temperature” means the temperature of a medium to be heat-exchanged with refrigerant in the heat-radiation side heat exchanger. Specifically, the external temperature is the indoor temperature when the heating operation is carried out, the outdoor air temperature when the outdoor heat exchanger functions as a radiator or the water temperature at the entrance of the water supply (stocking) heat exchanger when the water stocking operation is carried out.

The present invention is not limited to the above-described embodiment, and various modifications may be made without departing from the subject matter of the present invention.

In the foregoing description, at least one of the first and second throttling devices 5 and 9 is controlled so that the valve opening degree of one of the throttling devices is merely set to be “large” while the valve opening degree of the other throttling device is set to be “small” in accordance with whether the enthalpy of the refrigerant at the exit of the high-pressure gas cooler 3 is larger or not larger than the enthalpy of saturated liquid corresponding to the pressure of the intermediate-pressure portion of the compressor under one-stage expansion. In order to more finely control the amount of gas refrigerant to be supplied from the intermediate-pressure receiver to the intermediate-pressure portion of the compressor, the valve opening degrees of the first and second throttling devices 5 and 9 may be more finely controlled on the basis of the difference value between the enthalpy of the refrigerant at the exit of the high-pressure gas cooler 3 and the enthalpy of saturated liquid corresponding to the pressure of the intermediate-pressure portion of the compressor under one-stage expansion. 

1. A refrigerating machine comprising a two-stage compressor, a high-pressure gas cooler for cooling high-pressure gas refrigerant discharged from the two-stage compressor, a first throttling device for expanding the gas refrigerant from the high-pressure gas cooler, an intermediate-pressure receiver for adjusting a refrigerant circulating amount, a second throttling device for expanding the refrigerant from the intermediate-pressure receiver and an evaporator that are successively connected to one another to form a closed refrigerant circuit, further comprising: an intermediate-pressure refrigerant bypass circuit for bypassing gas refrigerant in the intermediate-pressure receiver to an intermediate-pressure portion of the two-stage compressor; a back flow preventing device that is provided to the intermediate-pressure refrigerant bypass circuit and prevents back flow of refrigerant from the two-stage compressor to the intermediate-pressure receiver; and a refrigerant-pressure control unit for controlling the pressure of the refrigerant in the intermediate-pressure receiver on the basis of the difference between a specific enthalpy of refrigerant discharged from the high-pressure gas cooler and a predetermined reference enthalpy.
 2. The refrigerating machine according to claim 1, wherein the predetermined reference enthalpy value is set to the enthalpy of saturated liquid corresponding to the pressure of the intermediate-pressure portion of the two-stage compressor.
 3. The refrigerating machine according to claim 1, wherein the first throttling device comprises a first expansion valve and the second throttling device comprises a second expansion valve.
 4. The refrigerating machine according to claim 3, wherein the refrigerant-pressure control unit controls the valve opening degrees of the first and second expansion values to thereby adjust the pressure of the refrigerant in the intermediate-pressure receiver.
 5. The refrigerating machine according to claim 1, wherein the refrigerant-pressure control unit controls at least one of the first throttling device and the second throttling device so that the pressure of the refrigerant in the intermediate-pressure receiver is lower than the pressure of the refrigerant in the intermediate-pressure portion of the two-stage compressor when the specific enthalpy of the refrigerant discharged from the high-pressure gas cooler is not larger than the predetermined reference enthalpy, and also so that the pressure of the refrigerant in the intermediate-pressure receiver is lower than the pressure of saturated liquid whose enthalpy is substantially equal to the specific enthalpy of the refrigerant discharged from the high-pressure gas cooler and also higher than the pressure of the refrigerant in the intermediate-pressure portion of the two-stage compressor when the specific enthalpy of the refrigerant discharged from the high-pressure gas cooler is larger than the predetermined reference enthalpy.
 6. The refrigerating machine according to claim 5, wherein the reference predetermined enthalpy value is set to the enthalpy of saturated liquid corresponding to the pressure of the intermediate-pressure portion of the two-stage compressor.
 7. A refrigerating machine comprising: an outdoor unit including a compressor having an intermediate pressure portion into which intermediate-pressure refrigerant having intermediate pressure between refrigerant pressure at a suction port of the compressor and refrigerant pressure at a discharge port of the compressor can be introduced, and an outdoor heat exchanger serving as a heat-source side heat exchanger, a plurality of indoor units each including an indoor heat exchanger serving as a using side heat exchanger, the plural indoor units carrying out one of cooling operation or heating operation at the same time or carrying out a mixing operation including cooling operation and heating operation at the same time; an inter-unit pipe for connecting the outdoor unit and each of the indoor units to each other, the inter-unit pipe comprising a high-pressure pipe connected to the refrigerant discharge pipe, a low-pressure pipe connected to the refrigerant suction pipe and an intermediate pressure pipe connected to the other end of the outdoor heat exchanger; an intermediate-pressure receiver that is disposed between the heat-source side heat exchanger and the using side heat exchanger and separates gas/liquid mixture refrigerant discharged from any one of the heat-source heat exchanger and the using side heat-exchanger into gas refrigerant and liquid refrigerant and supplying the gas refrigerant to the intermediate-pressure portion of the compressor; a back flow preventing device that is provided between the intermediate-pressure receiver and the intermediate-pressure portion of the compressor and prevents back flow of the gas refrigerant from the compressor to the intermediate-pressure receiver; and a refrigerant-pressure control unit for controlling the pressure of the refrigerant in the intermediate-pressure receiver on the basis of the difference between a predetermined reference enthalpy and a specific enthalpy of refrigerant discharged from any one of the heat-side heat exchanger and the using side heat exchanger.
 8. The refrigerating machine according to claim 7, wherein the refrigerant-pressure control unit comprises a first expansion valve disposed between the heat-source side heat exchanger and the intermediate-pressure receiver and a second expansion valve disposed between the intermediate-pressure receiver and the using side heat exchanger, and a controller for controlling at least one of the valve opening degrees of the first and second expansion valves on the basis of the difference between the predetermined reference enthalpy and the specific enthalpy of the refrigerant discharged from any one of the heat-side heat exchanger and the using side heat exchanger.
 9. The refrigerating machine according to claim 8, wherein when any one of the heat-side heat exchanger and the using side heat exchanger operates as a heat-radiation side heat exchanger, the controller controls at least one of the valve opening degrees of the first and second expansion valves so that the pressure of the refrigerant in the intermediate-pressure receiver is lower than the pressure of the refrigerant in the intermediate-pressure portion of the compressor when the specific enthalpy of the refrigerant discharged from the heat-radiation side heat exchanger is not larger than the predetermined reference enthalpy, and also so that the pressure of the refrigerant in the intermediate-pressure receiver is lower than the pressure of saturated liquid whose enthalpy is substantially equal to the specific enthalpy of the refrigerant discharged from the heat-radiation side heat exchanger and also higher than the pressure of the refrigerant in the intermediate-pressure portion of the compressor when the specific enthalpy of the refrigerant discharged from the heat-radiation side heat exchanger is larger than the predetermined reference enthalpy.
 10. The refrigerating machine according to claim 9, wherein the predetermined reference enthalpy value is set to the enthalpy of saturated liquid corresponding to the pressure of the intermediate-pressure portion of the compressor. 